Automatic four-speed transmission with interconnected clutches

ABSTRACT

An adaptive control system in an automatic four-speed transmission having a unique clutch and gear arrangement which eliminates the need for additional gearing and one over-running clutch, while requiring only one more friction element. The vehicle transmission assembly includes a transmission housing, a rotatable input member, an output member, and a gear assembly for changing the ratio of torque between the input member and the output member. Three clutch assemblies selectively couple the rotatable input member to predetermine gears of the gear assembly and two brake assemblies selectively couple predetermined gears of the gear assembly to the transmission housing. The gear assembly includes the first planetary gear set and a second planetary gear set. The first clutch assembly selectively couples the input member to the second sun gear, the second clutch assembly selectively couples the input member to the first planet carrier, the third clutch assembly selectively couples the input member to the first sun gear, the first brake assembly selectively couples the first sun gear to the transmission housing, and the second brake assembly selectively couples the first planet carrier and the second annulus gear to the transmission housing. An output member is connected to the second planet carrier to allow the output member and the second planet carrier to rotate at the same time.

This is a continuation of U.S. patent application Ser. No. 07/496,071, filed Feb. 27, 1990, abandoned which is a continuation of U.S. Ser. No. 07/187,751 filed Apr. 29, 1988, abandoned.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to an automatic transmission primarily intended for motor vehicle use, and more particularly, to a four-speed automatic transmission that is controlled electronically and hydraulically.

2. Description of Related Art

Generally speaking, land vehicles require three basic components. These components comprise a power plant (such as an internal combustion engine) a power train and wheels. The internal combustion engine produces force by the conversion of the chemical energy in a liquid fuel into the mechanical energy of motion (kinetic energy). The function of the power train is to transmit this resultant force to the wheels to provide movement of the vehicle.

The power train's main component is typically referred to as the "transmission". Engine torque and speed are converted in the transmission in accordance with the tractive-power demand of the vehicle. The vehicle's transmission is also capable of controlling the direction of rotation being applied to the wheels, so that the vehicle may be driven both forward and backward.

A conventional transmission includes a hydrodynamic torque converter to transfer engine torque from the engine crankshaft to a rotatable input member of the transmission through fluid-flow forces. The transmission also includes frictional units which couple the rotating input member to one or more members of a planetary gearset. Other frictional units, typically referred to as brakes, hold members of the planetary gearset stationary during flow of power. These frictional units are usually brake clutch assemblies or band brakes. The drive clutch assemblies can couple the rotating input member of the transmission to the desired elements of the planetary gearsets, while the brakes hold elements of these gearsets stationary. Such transmission systems also typically provide for one or more planetary gearsets in order to provide various ratios of torque and to ensure that the available torque and the respective tractive power demand are matched to each other.

Transmissions are generally referred to as manually actuated or automatic transmissions. Manual transmissions generally include mechanical mechanisms for coupling rotating gears to produce different ratio outputs to the drive wheels.

Automatic transmissions are designed to take automatic control of the frictional units, gear ratio selection and gear shifting. A thorough description of general automatic transmission design principals may be found in "Fundamentals of Automatic Transmissions and Transaxles." Chrysler Corporation Training Manual No. TM-508A. Additional descriptions of automatic transmissions may be found in U.S. Pat. No. 3,631,744 entitled "Hydromatic Transmission." issued Jan. 4, 1972 to Blomquist. et al., and U.S. Pat. No. 4,289,048, entitled "Lock-up System for Torque Converter," issued on Sept. 15, 1981 to Mikel, et al. Each of these patents is hereby incorporated by reference.

In general, the major components featured in such an automatic transmission are: a torque converter as above-mentioned; fluid pressure-operated multi-plate drive or brake clutches and/or brake bands which are connected to the individual elements of the planetary gearsets in order to perform gear shifts without interrupting the tractive power; one-way clutches in conjunction with the frictional units for optimization of power shifts; and transmission controls such as valves for applying and releasing elements to shift the gears (instant of shifting), for enabling power shifting, and for choosing the proper gear (shift point control), dependent on-shift program selection by the driver (selector lever), accelerator position, the engine condition and vehicle speed.

The control system of the automatic transmission is typically hydraulically operated through the use of several valves to direct and regulate the supply of pressure. This hydraulic pressure control will cause either the actuation or deactuation of the respective frictional units for effecting gear changes in the transmission. The valves used in the hydraulic control circuit typically comprise spring-biased spool valves, spring-biased accumulators and ball check valves. Since many of these valves rely upon springs to provide a predetermined amount of force, it will be appreciated that each transmission design represents a finely tuned arrangement of interdependent valve components. While this type of transmission control system has worked well over the years, it does have its limitations. For example, such hydraulically controlled transmissions are generally limited to one or a very small number of engines and vehicle designs. Therefore, considerable cost is incurred by an automobile manufacturer to design, test, build inventory and repair several different transmission units in order to provide an acceptable broad model line for consumers.

Additionally, it should be appreciated that such hydraulically controlled transmission systems cannot readily adjust themselves in the field to compensate for varying conditions such as normal wear on the components, temperature swings and changes in engine performance over time. While each transmission is designed to operate most efficiently within certain specific tolerances, typical hydraulic control systems are incapable of taking self corrective action on their own to maintain operation of the transmission at peak efficiency.

However, in recent years, a more advanced form of transmission control system has been proposed, which would offer the possibility of enabling the transmission to adapt itself to changing conditions. In this regard, U.S. Pat. No. 3,956,947 issued on May 18, 1976 to Leising, et al., which is hereby incorporated by reference sets forth a fundamental development in this field. Specifically this patent discloses an automatic transmission design which features an "adaptive" control system that includes electrically operated solenoid-actuated valves for controlling certain fluid pressures. In accordance with this electric/hydraulic control system, the automatic transmission would be "responsive" to an acceleration factor for controlling the output torque of the transmission during a shift from one ratio of rotation (between the input and output shafts of the transmission) to another. Specifically the operation of the solenoid-actuated valves would cause a rotational speed versus time curve of a sensed rotational component of the transmission to substantially follow along a predetermined path during shifting.

OBJECTS OF THE PRESENT INVENTION

It is one object of the present invention to provide a four-speed automatic transmission design which can be readily utilized in conjunction with a variety of engines and vehicle sizes and types, including vehicles presently using conventional, mechanical-hydraulic automatic transmission systems.

It is an additional object of the present invention to eliminate the need for a few certain elements (clutches, bands one-way clutches) which are normally required to provide acceptable shift quality.

It is a more specific object of the present invention to provide a unique clutch and gear arrangement which, relative to today's three-speed units, eliminates the need for additional gears of the gear assembly and one over-running (one-way) clutch, while requiring only one more friction element to increase the compactness and decrease the axial length of the transmission. Since the clutch and gear arrangement of the present invention eliminates the need for the low gear over-running clutch, the gearset pinions are not forced to spin at high speed when the vehicle is towed backwards. This reduces the need for lubrication and permits substantially unrestricted rearward or forward towing.

This application is one of several applications filed on the same date all commonly assigned and having similar. Specification and Drawings, these applications being identified below.

    __________________________________________________________________________     U.S. Ser. No.       Title                                                      __________________________________________________________________________     187,772, now U.S. Pat. No. 4,875,391                                                               AN ELECTRONICALLY-CONTROLLED,                                                  ADAPTIVE AUTOMATIC TRANSMISSION SYSTEM                     189,493, now U.S. Pat. No. 4,915,204                                                               PUSH/PULL CLUTCH APPLY PISTON OF AN                                            AUTOMATIC TRANSMISSION                                     187,781, abandoned in favor of                                                                     SHARED REACTION PLATES BETWEEN                             U.S. Ser. No. 463,870, now abandoned                                                               CLUTCH ASSEMBLIES IN AN AUTOMATIC TRANSMISSION             189,492, abandoned in favor of U.S. Ser.                                                           CLUTCH REACTION AND PRESSURE PLATES                        No. 432,311, now U.S. Pat. No. 4,958,753                                                           IN AN AUTOMATIC TRANSMISSION                               188,602 abandoned in favor of U.S. Ser.                                                            BLEEDER BALL CHECK VALVES IN AN                            No. 463,871, now U.S. Pat. No. 4,982,826                                                           AUTOMATIC TRANSMISSION                                     188,610 abandoned in favor of U.S. Ser.                                                            PRESSURE BALANCED PISTONS IN AN                            No. 488,419, now U.S. Pat. No. 4,964,506                                                           AUTOMATIC TRANSMISSION                                     189,494, now U.S. Pat. No. 5,027,934                                                               DOUBLE-ACTING SPRING IN AN AUTOMATIC TRANSMISSION          188,613, now U.S. Pat. No. 4,907,781                                                               PARK LOCKING MECHANISM FOR AN AUTOMATIC TRANSMISSION       187,770, now U.S. Pat. No. 4,887,491                                                               SOLENOID-ACTUATED VALVE ARRANGEMENT                                            OF AN AUTOMATIC TRANSMISSION SYSTEM                        187,796 abandoned in favor of U.S. Ser.                                                            RECIPROCATING VALVES IN A FLUID                            No. 532,149, now U.S. Pat. No. 5,031,656                                                           SYSTEM OF AN AUTOMATIC TRANSMISSION                        187,705, now U.S. Pat. No. 4,887,512                                                               VENT RESERVOIR IN A FLUID SYSTEM OF AN AUTOMATIC                               TRANSMISSION                                               188,592, now U.S. Pat. No. 4,936,166                                                               FLUID ACTUATED SWITCH VALVE IN AN AUTOMATIC                                    TRANSMISSION                                               188,598, now U.S. Pat. No. 4,893,652                                                               DIRECT-ACTING, NON-CLOSE CLEARANCE SOLENOID-ACTUATED                           VALVES                                                     188,618 abandoned in favor of U.S. Ser.                                                            NOISE CONTROL DEVICE FOR A SOLENOID-ACTUATED VALVE         No. 356,677                                                                    188,605, now U.S. Pat. No. 4,871,887                                                               FLUID ACTUATED PRESSURE SWITCH FOR AN AUTOMATIC                                TRANSMISSION                                               187,210, now U.S. Pat. No. 4,944,200                                                               METHOD OF APPLYING REVERSE GEAR OF AN                                          AUTOMATIC TRANSMISSION                                     187,672 abandoned in favor of U.S. Ser.                                                            TORQUE CONVERTER CONTROL VALVE IN A                        No. 356,667         FLUID SYSTEM OF AN AUTOMATIC TRANSMISSION                  187,120 abandoned in favor of U.S. Ser.                                                            CAM-CONTROLLED MANUAL VALVE IN AN AUTOMATIC                                    TRANSMISSION                                               No. 338,493, now U.S. Pat. No. 4,916,961                                       187,181, now U.S. Pat. No. 4,907,475                                                               FLUID SWITCHING MANUALLY BETWEEN VALVES IN AN                                  AUTOMATIC TRANSMISSION                                     187,704, now U.S. Pat. No. 4,975,845                                                               METHOD OF OPERATING AN ELECTRONIC AUTOMATIC                                    TRANSMISSION SYSTEM                                        188,020, now U.S. Pat. No. 4,935,872                                                               METHOD OF SHIFT SELECTION IN AN                                                ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM                   187,991 abandoned in favor of U.S. Ser.                                                            METHOD OF UNIVERSALLY ORGANIZING                           No. 320,883, now U.S. Pat. No. 4,936,167                                                           SHIFTS FOR AN ELECTRONIC AUTOMATIC TRANSMISSION                                SYSTEM                                                     188,603 abandoned in favor of U.S. Ser.                                                            METHOD OF DETERMINING AND                                  No. 470,439, now U.S. Pat. No. 4,993,527                                                           CONTROLLING THE LOCK-UP OF A TORQUE                                            CONVERTER IN AN ELECTRONIC AUTOMATIC TRANSMISSION                              SYSTEM                                                     188,617, now U.S. Pat. No. 4,965,728                                                               METHOD OF ADAPTIVELY IDLING AN ELECTRONIC AUTOMATIC                            TRANSMISSION SYSTEM                                        188,553             METHOD OF DETERMINING THE DRIVER SELECTED OPERATING                            MODE                                                                           OF AN AUTOMATIC TRANSMISSION SYSTEM                        188,615, now U.S. Pat. No. 4,965,735                                                               METHOD OF DETERMINING THE SHIFT LEVER POSITION OF AN                           ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM                   188,594, now U.S. Pat. No. 4,947,329                                                               METHOD OF DETERMINING THE ACCELERATION OF A TUR-                               BINE IN AN AUTOMATIC TRANSMISSION SYSTEM                   187,771, now U.S. Pat. No. 4,928,235                                                               METHOD OF DETERMINING THE FLUID TEMPERATURE OF AN                              ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM                   188,607 abandoned in favor of U.S. Ser.                                                            METHOD OF DETERMINING THE CONTINUITY                       No. 366,361, now U.S. Pat. No. 4,939,928                                                           OF SOLENOIDS IN AN ELECTRONIC AUTOMATIC TRANSMISSION                           SYSTEM                                                     189,579, now U.S. Pat. No. 4,975,844                                                               METHOD OF DETERMINING THE THROTTLE ANGLE POSITION FOR                          AN                                                                             ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM                   188,604, now U.S. Pat. No. 4,905,545                                                               METHOD OF CONTROLLING THE SPEED CHANGE OF A KICKDOWN                           SHIFT FOR AN ELECTRONIC AUTOMATIC TRANSMISSION                                 SYSTEM                                                     188,591, now U.S. Pat. No. 4,951,200                                                               METHOD OF CONTROLLING THE APPLY ELEMENT DURING A                               KICKDOWN                                                                       SHIFT FOR ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM         188,608             METHOD OF CALCULATING TORQUE FOR AN ELECTRONIC                                 AUTOMATIC                                                                      TRANSMISSION SYSTEM                                        187,150 abandoned in favor of U.S. Ser.                                                            METHOD OF LEARNING FOR ADAPTIVELY CONTROLLING AN                               ELECTRONIC                                                 No. 496,144, now U.S. Pat. No. 4,982,620                                                           AUTOMATIC TRANMISSION SYSTEM                               188,595 abandoned in favor of U.S. Ser.                                                            METHOD OF ACCUMULATOR CONTROL FOR A FRICTION ELEMENT                           IN AN                                                      No. 471,672, now abandoned                                                                         ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM                   188,599 abandoned in favor of U.S. Ser.                                                            METHOD OF ADAPTIVELY SCHEDULING A SHIFT FOR AN                                 ELECTRONIC                                                 326,492, now U.S. Pat. No. 4,938,102                                                               AUTOMATIC TRANMISSION SYSTEM                               188,601 abandoned in favor of U.S. Ser.                                                            METHOD OF SHIFT CONTROL DURING A COASTDOWN SHIFT FOR                           AN                                                         No. 529,697         ELECTRONIC AUTOMATIC TRANSMISSION SYSTEM                   188,620, now U.S. Pat. No. 4,969,098                                                               METHOD OF TORQUE PHASE SHIFT CONTROL FOR AN                                    ELECTRONIC                                                                     AUTOMATIC TRANSMISSION                                     188,596, now U.S. Pat. No. 4,951,205                                                               METHOD OF DIAGNOSTIC PROTECTION FOR AN ELECTRONIC                              AUTOMATIC TRANSMISSION SYSTEM                              188,597, now U.S. Pat. No. 4,991,097                                                               METHOD OF STALL TORQUE MANAGEMENT FOR AN ELECTRONIC                            AUTOMATIC TRANSMISSION SYSTEM                              188,606, now U.S. Pat. No. 4,968,999                                                               METHOD OF SHIFT TORQUE MANAGEMENT FOR AN ELECTRONIC                            AUTOMATIC TRANSMISSION SYSTEM                              188,616, now U.S. Pat. No. 4,998,200                                                               ELECTRONIC CONTROLLER FOR AN AUTOMATIC TRANSMISSION        188,600, now U.S. Pat. No. 4,959,786                                                               DUAL REGULATOR FOR REDUCING SYSTEM CURRENT DURING                              AT LEAST ONE MODE OF OPERATION                             188,619, now U.S. Pat. No. 4,992,951                                                               UTILIZATION OF A RESET OUTPUT OF A REGULATOR AS A                              SYSTEM LOW-VOLTAGE INHIBIT                                 188,593 abandoned in favor of U.S. Ser.                                                            THE USE OF DIODES IN AN INPUT CIRCUIT TO TAKE                                  ADVANTAGE                                                  No. 537,227, now U.S. Pat. No. 5,014,214                                                           OF AN ACTIVE PULL-DOWN NETWORK PROVIDED IN A DUAL                              REGULATOR                                                  188,609 abandoned in favor of U.S. Ser.                                                            SHUTDOWN RELAY DRIVER CIRCUIT                              No. 541,450, now U.S. Pat. No. 4,991,096                                       188,614, now U.S. Pat. No. 4,955,336                                                               CIRCUIT FOR DETERMINING THE CRANK POSITION OF AN                               IGNITION                                                                       SWITCH BY SENSING THE VOLTAGE ACROSS THE STARTER                               RELAY                                                                          CONTROL AND HOLDING AN ELECTRONIC DEVICE IN A RESET                            CONDITION IN RESPONSE THERETO                              188,612, now U.S. Pat. No. 4,901,561                                                               THROTTLE POSITION SENSOR DATA SHARED BETWEEN                                   CONTROLLER                                                                     WITH DISSIMILAR GROUNDS                                    188,611 abandoned in favor of U.S. Ser.                                                            NEUTRAL START SWITCH TO SENSE SHIFT LEVER POSITION         No. 492,269, now U.S. Pat. No. 4,998,450                                       188,981 abandoned in favor of U.S. Ser.                                                            OPEN LOOP CONTROL OF SOLENOID COIL DRIVER                  No. 492,235, now U.S. Pat. No. 4,980,793                                       __________________________________________________________________________

Commonly assigned application Ser. No. 07/187,772, filed Apr. 29, 1998, now U.S. Pat. No. 4,875,391 has been printed in its entirety. The Figures and the entire Specification of that application are specifically incorporated by reference. For a description of the above copending applications, reference is made to the above mentioned U.S. Pat. No. 4,875,391.

SUMMARY OF THE INVENTION

To achieve the foregoing objects, the present invention provides a comprehensive four-speed automatic transmission system. While this transmission system particularly features a fully adaptive electronic control system, numerous other important advances are incorporated into this unique transmission system, as will be described below in detail.

In addition to the advantages offered by the adaptive control system, the present invention achieves the combination of this control system with a unique four-speed transaxle structure which requires fewer components and is smaller than previous four-speed transmission systems. For example, the four-speed transmission system according to the present invention is capable of fitting into the space made available for a conventional three-speed transmission system.

Additionally, the four-speed transmission system features a simplified clutch and gear arrangement, which does not require over-running clutches.

BRIEF DESCRIPTION OF THE DRAWINGS

Other objects features and advantages of the present invention will become more fully apparent from the following detailed description of the preferred embodiment, the appended claims and in the accompanying drawings in which:

FIGS. 1A-E illustrate one physical embodiment of the transmission according to the present invention;

FIG. 1A is a perspective view, partially broken away of the transmission;

FIG. 1B is a sectional elevational view of one embodiment of a transmission constructed according to the present invention;

FIG. 1C is a partial sectional elevational view of one half of the transmission of FIG. 1B;

FIG. 1D is a partial sectional elevational view of the other half of the transmission of FIG. 1B; and

FIG. 1E is a schematic diagram of the transmission of FIGS. 1A and 1B;

FIGS. 2A and 2B illustrate two views of a bleeder ball check valve assembly according to the present invention;

FIG. 2A is a front elevational view of the reaction shaft support and bleeder ball check valve assembly; and

FIG. 2B is a sectional view of FIG. 2A;

FIGS. 3A-E illustrate the structure and operation of the double acting spring;

FIG. 3A is an elevational view of the structure of the double acting spring;

FIG. 3B is a sectional view taken along lines B--B of FIG. 3A;

FIGS. 3C is a partial sectional elevational view of the spring in its non-applied position;

FIG. 3D is a partial sectional elevational view of the spring while the overdrive clutch is being applied; and

FIG. 3E is a partial sectional elevational view of the spring while the reverse clutch is being applied.

DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring to FIGS. 1A and 1B, an automatic transmission 100 according to one embodiment of the present invention is shown. The transmission 100 is adapted to be used in a vehicle (not shown), such as an automobile. However, it should be appreciated that the principles of the present invention may be employed in other types of vehicles and devices. The transmission 100 includes a transmission housing or case 102 for enclosing the numerous subassemblies which make up the transmission 100, including a torque converter assembly 110, pump assembly 200. multi-clutch assembly 300 and gear assembly 500.

TORQUE CONVERTER ASSEMBLY STRUCTURE

The torque converter assembly 110 is operative to transmit power from a rotating crankshaft 114 of a prime mover such as an automobile engine (not shown) to the input member of the transmission 100. This power may then be subsequently transmitted to a drive unit 104 (partially shown) which is connected to one or more drive wheels (not shown) of the vehicle. The torque converter 110 is generally comprised of an impeller assembly 126 turbine assembly 128 and a stator assembly 130.

As illustrated in FIG. 1C. power is transmitted from the rotating crankshaft 114 of the engine to a front cover member 116 of the impeller assembly 126 through a rotatable plate member 118. Balance weights 119 are circumferentially spaced about the outer periphery of the front cover member 116. The plate member 118 is secured proximate its inner periphery to the crankshaft 114 by suitable fastening means such as bolts 120, and is likewise secured proximate its outer periphery to the front cover member 116 by suitable fastening means such as bolts 122. The front cover member 116 is secured, such as by welding at 124, to the impeller assembly 126 of the torque converter 110.

The impeller assembly 126 is fluidly connected in toroidal flow relationship in a known manner with the turbine assembly 128 and the stator assembly 130. The impeller assembly 126 comprises a plurality of circumferentially spaced impeller blades 132 connected to the inside of an impeller shell 134. The impeller shell 134 is secured, such as by welding at 136, to an impeller hub or pump drive shaft 138. The impeller hub 138 is drivingly engaged at its neck portion 140 to the positive displacement pump 200, from which fluid is supplied to the torque converter 110 in a manner to be described herein. An arcuate inner portion 142 of the impeller blade 132 is disposed about one half of a split torus ring 144 which reduces fluid turbulence within the torque converter 110. A thrust plate 146 is connected by tabs (not shown) on the inner surface of a slot 148 of the impeller hub 138 and disposed between the impeller assembly 126 and a stator thrust member 165.

The stator assembly 130 includes a plurality of circumferentially spaced stator vanes 150 which are connected at their inner end to a stator plate 152. The stator plate 152 is mounted on a one-way or over-running clutch assembly, generally indicated at 154. The over-running clutch assembly 154 permits rotation only in the direction of the impeller assembly 126. The over-running clutch assembly 154 comprises an over-running clutch cam 156 mounted about over-running clutch rollers 158 which in turn, travel about an over-running clutch race 160. The over-running clutch race 160 is splined at inner surface 162 to a stationary reaction shaft member 164. An annular thrust member 165 having retaining transverse flanges or tabs 167 is disposed between the stator plate 152 and the thrust plate 146.

The turbine assembly 128 includes a plurality of circumferentially spaced turbine blades 166 which ar connected to the inside of a turbine shell 168. The turbine shell 168 is secured by rivets 170 or the like to a turbine hub member 172. The turbine hub member 172 is drivingly connected, as by a spline connection 174, to a rotatable input member or shaft 176 to which the gear assembly 500 of the transmission 100 is drivingly engaged. A turbine hub seal 178 is disposed between the inside of the turbine hub member 172 and the input shaft 176 to prevent entry of fluid therebetween. A cover bushing 180 having grooves (not shown) for fluid flow therethrough supports the turbine hub member 172 in a cavity 182 of the front cover member 116. A thrust plate or washer 184 having grooves (not shown) for fluid flow therethrough is disposed between the turbine hub member 172 and the front cover member 116. An annular stepped member 185 having grooves (not shown) for fluid flow therethrough is disposed between the turbine hub member 172 and stator plate 152, as well as the over-running clutch race 160.

The torque converter 110 also includes a lock-up clutch assembly, generally indicated at 186, to prevent slip between the rotating crankshaft 114 of the engine and the turbine assembly 128 of the torque converter 110. The lock-up clutch assembly 186 includes an annular piston member 188 having an inner flange portion 190 disposed about the turbine hub member 172 of the turbine assembly 128. The piston member 188 has a plurality of circumferentially spaced inverted U-shaped outer flange portions 192 which are formed to engage corresponding slots 194 in a drive ring 196 that is welded to the turbine shell 168. The piston member 188 is slidingly and sealingly mounted for axial movement on the outer axial surface of the turbine hub member 172 through annular seal element 198. An annular disc-shaped frictional element or lock-up disc 199 is carried proximate the outer periphery of the front cover member 116 for engagement with a cooperating portion of the piston member 188.

PUMP ASSEMBLY STRUCTURE

The fixed positive displacement pump assembly 200 includes a pump housing 202 secured proximate its inner periphery to a reaction shaft support 204 by suitable fastening means such as bolts 206. The pump housing 202 is likewise secured proximate its outer periphery to the transmission case 102 by suitable fastening means such as bolts 208. The reaction shaft support 204 is secured, such as by press fitting with splines at 210, to the reaction shaft member 164. The impeller hub or pump drive shaft 138 is supported in the pump housing 202 through a bushing member 212. A seal ring assembly 214 is disposed about the impeller hub or pump drive shaft 138 in a bore or recess 216 at one end of the pump housing 202 to prevent fluid from exiting the end of the pump housing 202. An outer gear or rotor 218 with internal teeth (not shown) operates within a bore 220 of the pump housing 202. An inner gear or rotor 222 having external teeth (not shown), cooperative with the teeth of the outer rotor 218, is disposed within the outer rotor 218. As illustrated in FIGS. 1C and 1D, sealing means, such as seal rings 224, 226 and 228 are axially spaced between the input shaft 176 and reaction shaft support 204. The reaction shaft support 204 includes a fluid passage 230 to allow fluid to flow to the torque converter 110 in a manner to described herein.

MULTI-CLUTCH ASSEMBLY STRUCTURE

During the flow of power through the transmission 100, the multi-clutch assembly 300 provides a means for application and release of two separate members to and from each other. In other words, the multi-clutch assembly 300 is the means by which the gears within the transmission are selectively engaged and disengaged from either the crankshaft 114 of the prime mover or the transmission case 102. Near the input side of the transmission 100, the multi-clutch assembly 300 includes an underdrive clutch 302 (applied in first, second and third gears). overdrive clutch 304 (applied in third and fourth gears) and a reverse clutch 306 (applied in reverse gear) assemblies. Near the output side of the transmission 100, the multi-clutch assembly 300 includes a two/four shift clutch brake assembly 308 (applied in second and fourth gears), and a low/reverse clutch brake assembly 310 (applied in first and reverse gears).

As illustrated in FIGS. 1C and 1D, an input clutch retainer hub 312 is provided to house the input clutch assemblies 302. 304 and 306. The input clutch retainer hub 312 has a generally axially extending shoulder portion 313 and a generally axially extending portion 314. A plurality of spaced seal rings 315 are disposed in corresponding annular grooves 316 which are formed along the reaction shaft support 204. The input clutch retainer hub 312 is also splined at 317 to the input shaft 176. A thrust bearing 318 is disposed axially between one end of the reaction shaft support 204 and the axially extending portion 314 of the input clutch retainer hub 312. The input clutch retainer hub 312 has teeth 319 at its outer periphery. A turbine speed sensor 320 threadably engages a bore 322 in the transmission case 102 and has one end 324 disposed or spaced radially just above the teeth 319 of the input clutch retainer hub 312. The turbine speed sensor 320 is used to monitor or sense the revolution rate of the turbine assembly 128 by counting the teeth 319 passing thereby in relation to time. Preferably, a passive type speed sensor is used for the turbine speed sensor 320.

An input clutch retainer 326 has a hub portion 328 disposed about and drivingly connected to, as by a spline connection 330, to the axially extending portion 314 of the input clutch retainer hub 312. Sealing means, such as sealing rings 332 and 334, are disposed in corresponding grooves of the input clutch hub retainer 312 between the hub portion 328 and the axially extending portion 314 of the input clutch retainer hub 312. A tapered snap ring 336 is disposed in a groove 338 of the input clutch retainer hub 312 to prevent axial movement of the input clutch retainer 326 toward the gear assembly 500. The input clutch retainer 326 includes an axially extending flange 340 forming a cylinder. A plurality of circumferentially spaced clutch retainer fingers 341 extend radially inwardly from the flange 340 to which the clutch plates, which will be described herein, are mounted.

As illustrated in FIG. 1D. the underdrive clutch assembly 302 comprises a plurality of axially spaced annular clutch plates 342 and a plurality of axially spaced annular clutch discs 344. The clutch discs 344 are alternated between the clutch plates 342 and when the clutch assembly 302 is not applied, these plates and discs are free to move or rotate relative to each other. The clutch plates 342 have splines (not shown) on their outer diameter and mount in grooves 346 of the clutch retainer fingers 341 which are inside the input clutch retainer 326. The clutch discs 344 have internal splines (not shown) and are lined with a friction material 347. The clutch discs 344 are mounted in grooves 348 in an underdrive clutch hub 350. The underdrive clutch hub 350 is integral with a rotatable underdrive gear shaft 352 of the gear assembly 500. A thrust bearing 353 is disposed axially between the axially extending portion 314 of the input clutch retainer hub 312 and underdrive clutch hub 350.

The overdrive clutch assembly 304 comprises a plurality of axially spaced annular clutch plates 354 and a plurality of axially spaced annular clutch discs 356. The clutch plates 354 and clutch discs 356 are similar to those of the underdrive clutch assembly 302. Clutch discs 356 are disposed in splines formed in an overdrive clutch hub 358 which is supported by bushings 360 and 361 about the gear shaft 352. Thrust members 362 and 363 are disposed axially between the underdrive clutch hub 350 and overdrive clutch hub 358. The thrust members 362 and 363 are similar to the thrust member 165. An annular reaction plate 364 is secured to the inside of the input clutch retainer 326 axially between the underdrive and overdrive clutch plates and discs 342. 344 354 and 356, respectively. The reaction plate 364 is shared by the underdrive 302 and overdrive 304 clutch assemblies. Annular snap rings 366 and 368 are disposed on the sides of the reaction plate 364. Snap ring 368 is a tapered ring, restraining reaction plate 364 from axial movement.

The reverse clutch assembly 306 comprises at least one annular clutch plate 370 and a plurality of axially spaced annular clutch discs 372. The reverse clutch plate 370 and clutch discs 372 are similar to those of the underdrive clutch assembly 302. The reverse clutch discs 372 are mounted in splines 373 of a reverse clutch hub 374. The reverse clutch hub 374 is supported by bushings 376 and 378 about one end of the overdrive clutch hub 358 A thrust member 379 is disposed axially between the overdrive clutch hub 358 and reverse clutch hub 379. The thrust member 379 is similar to the thrust member 165. An annular reaction plate 380 is mounted about one end of the flange 340 of the input clutch retainer 326 on one side of the reverse clutch plate 370 and discs 372. Selective snap rings 384 secure the reaction plate 380 from axial movement along the input clutch retainer 326.

To apply the overdrive clutch assembly 304 and reverse clutch assembly 306, a fluid actuating device such as a first hydraulic piston 386 has an axially extending projection 388 which operates in a bore or recess 390 of the input clutch retainer hub 312. The inner diameter of the projection 388 has a groove 392 provided for a snap ring 394, while the recess 390 of the input clutch retainer hub 312 has a groove 396 for a sealing means such as a synthetic rubber seal ring 398. The first hydraulic piston 386 is slidingly and sealingly mounted for axial movement on the outer diameter of the hub portion 328 of the input clutch retainer 326 through sealing means 400 at its inner periphery and near the outer periphery of the input clutch retainer 326 through sealing means 402. A double-acting spring means such as a Belleville like spring 404 is disposed between the first hydraulic piston 386 and the input clutch retainer hub 312 to bias or return the first hydraulic piston 386 to its non-displaced or non-applied position shown in the figure. The double-acting spring 404 has a conical shape with fingers 405 and is formed with a linear slope such that its inner and outer diameters do not lie in the same cross-sectional plane. The double-acting spring 404 will be discussed more in detail under the section heading "DOUBLE-ACTING SPRING".

The first hydraulic piston 386 includes an axially extending cylinder portion 406 which has an annular pressure plate member 408 secured at one end thereof by waved snap ring 409a and snap ring 409b. A pressure plate member 408 is interposed between the overdrive clutch assembly 304 and the reverse clutch assembly 306 to engage the clutch plates 354. 370 and discs 356, 372, respectively. Hence, the single pressure plate member 408 is shared by the overdrive clutch 304 and reverse clutch 306 assemblies.

To engage or disengage the underdrive clutch assembly 302, a second hydraulic piston 410 operates in a recess 412 of the input clutch retainer 326. The smooth outer diameter of the hub portion 314 of the input clutch retainer hub 312 has a groove 414 provided with a sealing means such as a synthetic rubber inner seal ring 416, while the outer periphery of recess 412 has a groove 418 for an outer seal ring 420. The second hydraulic piston 410 has one end 422 abutting the clutch plates 342 of the underdrive clutch assembly 302. An annular conically shaped spring retainer member 424 is abuttingly mounted against a snap ring 426. The snap ring 426 is disposed in a groove 428 formed in the axially extending portion 314 of the input clutch retainer hub 312. The other end of the spring retainer member 424 is in sealing engagement with the second hydraulic piston 410 through sealing means 430. The spring retainer member 424 is filled with fluid fed through an orifice (not shown) in the second hydraulic piston 410 from a passage (not shown) in the input clutch retainer hub 312 to provide the pressure balance for the second hydraulic piston 410. The excess fluid is allowed to leak past the snap ring 426 to cool the underdrive clutch assembly 302. A spring means such as a coiled spring 432 is disposed between the spring retainer member 424 and the second hydraulic piston 410 to bias or return the second hydraulic piston 410 to its original position shown in the figure when not applied.

At the output end of the transmission 100, the transmission case 102 houses the output or brake clutch assemblies such as the two/four shift brake assembly 308 and the low/reverse brake assembly 310. The two/four shift brake assembly 308 comprises a plurality of axially spaced annular clutch plates 434 and a plurality of axially spaced annular clutch discs 436. The clutch plates 434 and clutch discs 436 are similar to those of the underdrive clutch assembly 302. The clutch plates 434 are mounted in splines 438 of circumferentially spaced and radially inwardly extending case clutch fingers 439 inside the transmission case 102. The clutch discs 436 are mounted in splines 440 formed in an axially extending flange 442 of the reverse clutch hub 374. A spring means such as a Belleville like spring 444, similar to spring 404, is mounted inside the transmission case 102 on one side of the two/four shift brake assembly 308. An annular reaction plate 445 is mounted on the other side of the two/four shift brake assembly 308 and between the two/four shift brake assembly 308 and the low/reverse brake assembly 310. The reaction plate 445 is shared by the two/four shift brake 308 and low/reverse brake 310 assemblies. Snap rings 446 and 447 are mounted in the transmission case 102 on the sides of the reaction plate 445 to lock it in place. Snap ring 446 is a tapered ring, restraining reaction plate 445 from axial movement.

To apply the two/four shift brake assembly 308, a third hydraulic piston 448 operates in a cavity 450 formed by an annular piston housing 452. The piston housing 452 is secured to the transmission case 102 by suitable fastening means (not shown). The smooth diameter of the third hydraulic piston 448 has a groove 454 formed in its outer periphery for an outer seal ring 456 and a groove 458 formed in its inner periphery for an inner seal ring 460. A snap ring 462 is disposed in a groove 464 in the transmission case 102 to prevent axial movement of the piston housing 452.

The low/reverse brake assembly 310 comprises a plurality of axially spaced annular clutch plates 466 and a plurality of axially spaced annular clutch discs 468. The clutch plates 466 and clutch discs 468 are similar to those of the underdrive clutch assembly 302. The clutch plates 466 are mounted in splines 470 of the case clutch fingers 439 inside the transmission case 102. The clutch discs 468 are mounted in splines 472 of the outer periphery of an annulus gear 542 of the gear assembly 500 to be described further herein.

To apply the low/reverse brake assembly 310, a fourth hydraulic piston 474 operates in a cavity 476 formed by an annular piston housing 478. The piston housing 478 is disposed in an annular recess 480 of the transmission case 102 and secured by suitable fastening means, such as bolts 481, to the transmission case 102. The smooth diameter of the fourth hydraulic piston 474 has a groove 482 formed in its outer periphery for an outer seal ring 484 and a groove 486 formed in its inner periphery for an inner seal ring 488. A spring means such as a Belleville like spring 490, similar to spring 404, is disposed between the fourth hydraulic piston 474 and the gear assembly 500 to bias or return the fourth hydraulic piston 474 to its original position when not applied as shown in the figure. A snap ring 492 retains one end of the spring 490 to the transmission case 102.

GEAR ASSEMBLY STRUCTURE

During the flow of power, the gear assembly 500 changes the ratio of torque between an input member, such as input shaft 176, and an output member, such as output gear 534 which will be further described herein. The gear assembly 500 comprises a front or first planetary gear set, generally indicated at 502, and an axially spaced rear or second planetary gear set generally indicated at 504. The first planetary gear set 502 includes a first sun gear 506 at its center. The first sun gear 506 is connected to the reverse clutch hub 374 at its inner periphery and is supported upon bushings 376 and 378. A first planet carrier 508 is disposed about the first sun gear 506. The first planet carrier 508 includes a plurality of circumferentially spaced first pinion gears 510 mounted about shafts 512 connected to the first planet carrier 508. The first planet carrier 508 includes an inner portion 514 splined at 516 to the overdrive clutch hub 358. A thrust bearing 517 is disposed axially between one end of the first sun gear 506 and inner portion 514 of the first planet carrier 508 The first planet carrier 508 also includes an axially extending outer portion 518 forming a cylinder about the first planetary gear set 502. A first annulus gear 519 is disposed about the first planet carrier 508 and engages the first pinion gears 510.

The rear or second planetary gear set 504 includes a second sun gear 520 at its center which is splined at 522 to the gear shaft 352. A thrust bearing 523 is axially disposed between one end of the inner portion 514 of the first planet carrier 508 and the second sun gear 520. A second planet carrier 524 is disposed about the second sun gear 520. The second planet carrier 524 includes a plurality of circumferentially spaced second pinion gears 526 mounted about shafts 528 connected to the second planet carrier 524. The second planet carrier 524 includes an inner portion 530 splined at 532 to a rotatable output gear 534 which acts as the output member of the transmission 100. The inner portion 530 of the second planet carrier 524 is supported by a bushing 536 disposed about the second sun gear 520. A thrust bearing 537 is disposed axially between the second sun gear 520 and second planet carrier 524. A tapered roller bearing assembly 538 supports the inner portion 530 of the second planet carrier 524 within the transmission case 102.

The second planet carrier 524 also includes an outer portion 540 connected to the first annulus gear 519. The second annulus gear 542 is disposed about the second planet carrier 524 and engages the second pinion gears 526. The second annulus gear 542 is connected to the outer portion 518 of the first planet carrier 508.

The second planet carrier 524 includes teeth 544 at its outer periphery of the outer portion 540. An output speed sensor 546 threadably engages a bore 548 in the transmission case 102 and has one end 550 disposed or radially spaced just above the teeth 544 of the second planet carrier 524. The output speed sensor 546 is used to monitor or sense the revolution rate (per minute) of the second planet carrier 524 by counting or sensing the teeth 544 passing thereby relative to time. The output speed sensor 546 is similar to the turbine speed sensor 320. It should also be noted that other suitable speed sensors could be used inside or after the transmission 100 to provide an output speed signal to the transmission's controller 3010.

The output gear 534 is secured to the second planet carrier 524 by suitable fastening means such as a bolt 552. The output gear 534 is supported by a tapered roller bearing assembly 554 within the transmission case 102. A rear cover plate member 556 is connected by suitable fastening means (not shown) to the rear or output end of the transmission case 102 to enclose the output gear 534 and the transfer gear (not shown).

To visualize and understand how power is transmitted from the rotating crankshaft 114 of the engine to the output gear 534 of the transmission 100, the operation of the assemblies described above will now be discussed in connection with FIGS. 1C, 1D and 1E.

OPERATION OF THE TORQUE CONVERTER

Rotation of the crankshaft 114 of the engine causes the front cover member 116 to rotate with it due to the connection between bolts 120, plate member 118 and bolts 122. Since the front cover member 116 is welded at 124 to the impeller shell 134 of the impeller assembly 126, the impeller assembly 126 also rotates with the crankshaft 114. The fluid within the impeller assembly 126 is set into motion by the rotation of the impeller assembly 126 and by the fluid pressure from the pump assembly 200. The impeller blades 132 start to carry the fluid around with them. As the fluid is spun around by the impeller blades 132, it is thrown outward by centrifugal force and into the turbine assembly 128 at an angle. The fluid strikes the turbine blades 166 of the turbine assembly 128, thus imparting torque, or turning effort to the turbine assembly 128 and causing the turbine shell 168 and the turbine assembly 128 to rotate. Since the turbine shell 168 is connected to the turbine hub 172 through rivets 170 and the turbine hub 172 is splined at 174 to the input shaft 176, the input shaft 176 is caused to rotate. As engine speed is increased, the force of the fluid striking the turbine blades 166 is also increased. Thus, torque is imparted to the input shaft 176 of the transmission 100 via the turbine assembly 128.

In the torque converter 110 the stator assembly 130 redirects the fluid flow so that the turbine blades 166 will have more force exerted upon them during a torque multiplication stage. During torque multiplication, the over-running clutch assembly 154 in the stator assembly 130 is locked in a known manner so that the stator assembly 130 will remain stationary. As the fluid passes from the turbine assembly 128 to the impeller assembly 126, the stator blades 150 of the stator assembly 130 "push" the fluid against the impeller blades 132 so that a greater entry angle is imparted to the turbine blades 166, resulting in a greater force on the blades 166 and increasing the torque to the input shaft 176 of the transmission 100.

The over-running clutch assembly 154 also permits the stator assembly 130 to rotate only in the same direction as the impeller assembly 126. The over-running clutch assembly 154 resists torque in one direction for the purpose of making the stator plate 152 and stator vanes 150 stationary. This is accomplished by the clutch rollers 158 engaging radially narrowing recesses (not shown) in the over-running clutch cam 156 to cause the over-running clutch cam 156, rollers 158 and race 160 to form a single unit. Since the over-running clutch race 160 is splined at 162 to the reaction shaft 164 which, in turn, is welded at 210 to the reaction shaft support 204 which cannot rotate, the over-running clutch cam 156, rollers 158 and race 160 remain stationary, resulting in the stator plate 152 and vanes 150 remaining stationary. The over-running clutch assembly 154 allows the stator plate 152 and vanes 150 to rotate freely in the opposite direction when their function as a reaction member is not desired because the rollers 158 do not engage the recesses, resulting in the over-running clutch cam 156 rotating freely about the clutch race 160.

OPERATION OF TORQUE CONVERTER LOCK-UP

The lock-up function of the torque converter 110 will now be described. Fluid flows through the center passage 175 of the input shaft 176 into the cavity 182 of the front cover member 116. The turbine hub seal 178 prevents leakage of the fluid back around the input shaft 176. The fluid in cavity 182 flows through slots (not shown) in the front cover bushing 180 and the thrust washer 184 and against the lock-up piston 188. The fluid pushes the portion 192 of the lock-up piston 188 off the friction disc 199, resulting in non-lock-up operation. At the same time, fluid from the pump assembly 200 flows through passage 230 in the reaction shaft support 204 and between the input shaft 176 and reaction shaft member 164. This fluid flows through slots (not shown) in the stepped member 185 and into the turbine 128, stator 130 and impeller 126 assemblies of the torque converter 110. Fluid also flows from these assemblies 126, 128 and 130 between the lock-up piston 188 and the turbine shell 168. Hence, during normal torque converter operation, fluid flow is acting on the opposite side of the lock-up piston 188, attempting to apply the lock-up piston 188. When the input shaft fluid is vented, the torque converter fluid pushes the lock-up piston 188 against the front cover member 116 with the friction disc 199 sandwiched between the two elements. Engine torque can then go through the front cover member 116 to the lock-up piston 188 and, in turn, to drive ring 196 and turbine shell 168.

As will be appreciated, lock-up of the torque converter 110 is desirable to reduce or eliminate rotational speed difference or "slip" between the crankshaft 114 of the engine and the input shaft 176 of the transmission 100 Lock-up of the torque converter 110 may be partial or full lockup. Partial lockup will reduce slip to predetermined value. Full lockup will eliminate slip or reduce it to a zero value. Lockup of the torque converter 110 may occur in second, third and fourth gears.

OPERATION OF PUMP

The general operation of the pump assembly 200 will now be described. Specific fluid flow from the pump 200 to various assemblies in the transmission 100 will be described in other sections herein.

The pump 200 creates flow and applies force to the fluid. As described previously, the impeller shell 134 is welded at 136 to the impeller hub 138 which acts as the pump drive shaft of the pump assembly 200. Rotation of the impeller shell 134 results in rotation of the impeller hub 138 Thus, the external source of power for the pump 200 is the engine.

In the pump assembly 200, both rotor members 218 and 222 rotate together. The inner rotor 222 is splined at 140 to the impeller hub 138 and, therefore, rotates as the impeller hub 138 rotates. As the inner rotor 222 rotates or drives the outer rotor 218, a space (not shown) between the rotors 218. 222 increases as the rotor teeth separate and pass an outlet port (not shown).

In the pump assembly 200, a crescent shaped protrusion (not shown) of the pump housing 202 divides the rotors 218 and 222. Fluid is trapped between the protrusion and the rotor teeth as it is carried to the outlet port for further use in a manner to be described in other sections herein.

OPERATION OF THE CLUTCHES

As described previously, the input shaft 176 of the transmission 100 is rotating due to torque being transferred from the rotating crankshaft 114 of the engine and through the torque converter 110 to the input shaft 176. The input clutch retainer hub 312 also rotates with the input shaft 176 due to its spline connection 317 with the input shaft 176. The input clutch retainer 326 and clutch plates 342. 354 and 370 also rotate with the input shaft 176 due to the spline connection 330 of the input clutch retainer 326 to the input clutch retainer hub 312 and spline connection of clutch plates 342, 354 and 370 to the input clutch retainer 326.

To apply the underdrive clutch assembly 308, hydraulic pressure from fluid entering between the input clutch retainer 326 and second hydraulic piston 410 moves the second hydraulic piston 410 axially, thereby compressing the spring 432. The second hydraulic piston 410 forces the rotating clutch plates 342 and momentarily stationary discs 344 of the underdrive clutch assembly 302 together and produces frictional force between the clutch plates 342 and discs 344. Because the input clutch retainer 326 and underdrive clutch plates 342 are rotating, the frictional force causes the underdrive clutch discs 344 and hub 350 to rotate, in turn, rotating gear shaft 352 of the gear assembly 500. When the hydraulic fluid to the underdrive clutch assembly 302 is vented, the compressed spring 432 applies a force to the second hydraulic piston 410, thereby returning the second hydraulic piston 410 to its non-applied position as shown in the figure.

To apply the overdrive clutch assembly 304, hydraulic pressure from fluid entering between the first hydraulic piston 386 and the input clutch retainer 326 moves or pulls the first hydraulic piston 386 axially, thereby deflecting axially the spring 404. The pressure plate member 408 of the first hydraulic piston 386 forces the clutch plates 354 and discs 356 of the overdrive clutch assembly 304 together against the reaction plate 364 and produces a frictional force between them. Because the input clutch retainer 326 and overdrive clutch plates 354 are rotating the frictional force causes the overdrive clutch discs 356 and overdrive clutch hub 358 to rotate, in turn, rotating the first planet carrier 508 and second annulus gear 542. When the hydraulic fluid to the overdrive clutch assembly 304 or first hydraulic piston 386 is vented, the deflected spring 404 applies a force to the first hydraulic piston 386, thereby returning the first hydraulic piston 386 to its non-applied position as shown in the figure.

To apply the reverse clutch assembly 306, hydraulic pressure from fluid entering between the first hydraulic piston 386 and input clutch retainer hub 312 moves or pushes the first hydraulic piston 386 axially, thereby deflecting the spring 404. The pressure plate member 408 of the first hydraulic piston 386 forces the clutch plate 370 and discs 372 of the reverse clutch assembly 306 together against the reaction plate 380 and produces a frictional force between them. Because the input clutch retainer 326 and reverse clutch plate 370 are rotating, the frictional force causes the reverse clutch discs 372 and reverse clutch hub 374 to rotate, in turn, rotating the first sun gear 506. When the hydraulic fluid to the reverse clutch assembly 306 or first hydraulic piston 386 is vented, the deflected spring 404 applies a force to the first hydraulic piston 386, thereby returning the first hydraulic piston 386 to its non-applied position as shown in the figure.

At the output end of the transmission 100, the two/four shift brake 308 and low/reverse brake 310 assemblies are used to hold a particular gear element of the gear assembly 500 against rotation by coupling it to the relatively stationary transmission case 102. To apply the two/four shift brake assembly 308, hydraulic pressure from fluid entering between the third hydraulic piston housing 452 and the third hydraulic piston 448 moves the third hydraulic piston 448 axially, thereby deflecting the spring 444. The third hydraulic piston 448 forces the brake plates 434 and discs 436 of the two/four shift clutch assembly 308 together against the reaction plate 445 and produces a frictional force between them Because the two/four clutch plates 434 do not rotate or are stationary, as they are connected to the transmission case 102, the frictional force holds the two/four clutch discs 436 stationary, in turn, holding the flange 442, reverse hub member 374 and first sun gear 506 stationary. When the hydraulic fluid to the two/four shift brake assembly 308 or third hydraulic piston 448 is vented, the deflected spring 444 applies a force to the third hydraulic piston 448, thereby returning the third hydraulic piston 448 to its non-applied position as shown in the figure.

To apply the low/reverse brake assembly 310, hydraulic pressure from fluid entering between the fourth hydraulic piston housing 476 and the fourth hydraulic piston 474 moves the fourth hydraulic piston 474 axially, thereby deflecting the spring 490. The fourth hydraulic piston 474 forces the brake plates 466 and discs 468 of the low/reverse clutch assembly 310 together against reaction plate 445 and produces a frictional force between them. Because the low/reverse clutch plates 466 are stationary, as they are connected to the transmission case 102, the frictional force holds the low/reverse clutch discs 468 stationary, in turn, holding the second annulus gear 542 and first planet carrier 508 stationary. When the hydraulic fluid to the low/reverse clutch assembly 474 or fourth hydraulic piston 474 is vented, the deflected spring 490 applies a force to the fourth hydraulic piston 474, thereby returning the fourth hydraulic piston 474 to its non-applied position as shown in the figure.

OPERATION OF PLANETARY GEARS

In the neutral N or park P modes of transmission operation, the input shaft 176 (which is attached to the turbine assembly 128) freely rotates with the engine crankshaft 114. Since the input clutch retainer hub 312 is also attached to the input shaft 176, the input clutch retainer hub 312 rotates, in turn, causing the input clutch retainer 326 and clutch plates 342 354 and 370 to freely rotate with the engine crankshaft 114.

When the transmission 100 is desired to operate in first gear, the underdrive clutch assembly 302 and low/reverse clutch assembly 310 are applied. Hydraulic fluid moves the second hydraulic piston 410 axially away from the torque converter 110 to engage the clutch plates 342 and friction discs 344 of the underdrive clutch assembly 302. This engagement causes the underdrive clutch hub 350 to rotate which, in turn, rotates the gear shaft 352. Because the second sun gear 520 is splined at 522 to the gear shaft 352, rotation of the gear shaft 352 causes the second sun gear 520 to rotate. As the low/reverse clutch assembly 310 is applied by the engagement of the low/reverse clutch plates 466 with the discs 468, the second annulus gear 542 is held stationary. Since the second annulus gear 542 is connected to the first planet carrier 508 the first planet carrier 508 is held stationary. As a result, rotation of the second sun gear 520 causes rotation of the second pinion gears 528 and the second planet carrier 524. Because the output gear 534 is splined at 532 to the second planet carrier 524, rotation of the second planet carrier 524 causes the output gear 534 to rotate. Since the second planet carrier 524 rotates, the first annulus gear 519 also rotates, causing the first pinion gears 510 and first sun gear 506 to freely rotate in first gear. The output gear 534 then transfers the torque from the second planetary carrier 524 to the transfer gear (not shown).

When the transmission 100 is desired to operate in second gear, the underdrive brake assembly 302 and the two/four shift clutch assembly 308 are applied. Once again, the underdrive clutch hub 350 rotates as described above which, in turn, rotates the gear shaft 352. Rotation of gear shaft 352 causes the second sun gear 520 to rotate. As the two/four shift brake assembly 308 is applied by engagement of the two/four shift clutch plates 434 with the discs 436, the flange 442, reverse clutch hub 374 and first sun gear 506 are held stationary. Because the transmission 100 has been operating in first gear, the first annulus gear 519 and second planet carrier 524 have been rotating at output speed. Also, the first sun gear 506 has been rotating freely. By holding the first sun gear 506 stationary, the first pinion gears 510 and first planet carrier 508 increase in speed. As a result, the first annulus gear 519, second planet carrier 524 and the output gear 534 rotate at a greater r.p.m. than first gear.

When the transmission 100 is desired to operate in third gear, the underdrive clutch assembly 302 and the overdrive clutch assembly 304 are applied. Once again, engagement of the underdrive clutch assembly 302 causes the second sun gear 520 to rotate as previously described. As the overdrive clutch assembly 304 is applied by engagement of the clutch plates 354 and discs 356 of the overdrive clutch assembly 304, the overdrive clutch hub 358 rotates, in turn, rotating the first planet carrier 508 due to the spline connection at 516. Since the first planet carrier 508 rotates, the first pinion gears 510, first sun gear 506 and second annulus gear 542 also rotate. As a result, the second pinion gears 526 of the second planet carrier 524 rotate, causing the second planet carrier 524 to rotate which, in turn, rotates the output gear 534 at input speed or a higher r.p.m. than second gear.

When the transmission 100 is desired to operate in fourth gear, the overdrive brake assembly 304 and two/four shift clutch assembly 308 are applied. Application of the overdrive clutch assembly 304 causes the overdrive clutch hub 358 to rotate, as previously described. Rotation of the overdrive clutch hub 358 causes the first planet carrier 508 and second annulus gear 542 to rotate. Application of the two/four shift brake assembly 308 causes the flange 442, reverse clutch hub 374 and first sun gear 506 to be held stationary as previously described. As a result, rotation of the first planet carrier 508 causes the first pinion gears 510, first annulus gear 519 and second annulus gear 542 to rotate. Rotation of the first and second annulus gears 519 and 542, respectively, causes the second pinion gears 526 and second planet carrier 524 to rotate which, in turn, rotates the output gear 534 at a greater r.p.m. than third gear.

When the transmission 100 is desired to operate in reverse gear, the reverse brake assembly 306 and low/reverse clutch assembly 310 are applied. The reverse clutch assembly 306 is applied by engagement of the reverse clutch plate 370 and discs 372. This engagement causes the reverse clutch hub 374 to rotate which, in turn, rotates the first sun gear 506. Application of the low/reverse brake assembly 310 causes the first planet carrier 508 and the second annulus gear 542 to be held stationary as previously described. As a result, the first sun gear 506 rotates the first pinion gears 510 which, in turn, rotate the first annulus gear 519 backwards. Rotation of the first annulus gear 519 causes the second planet carrier 524 and second pinion gears 526 to rotate which, in turn, causes rotation of the output gear 534 in a direction opposite to the other gear positions. Rotation of the second pinion gears 526 also causes the second sun gear 520 to rotate freely.

CLUTCH REACTION AND APPLY PLATES

Referring to FIG. 1D. the reaction plate 380 and pressure plate member 408 are shown. The present invention features web means such as an annular web 380a spaced radially at the outer periphery of the reaction plate 380 and connected to the reaction plate 380 at least one location circumferentially and an annular web 408a spaced radially at the outer periphery of the pressure plate member 408 and connected to the pressure plate member 408 at least one location circumferentially. The webs 380a and 408a are an efficient means of increasing axial rigidity to restrict clutch deflection. The reaction plate 380 and pressure plate member 408 resist loads producing a stress pattern like that found in a Belleville spring (i.e. producing a family of moments along the radial direction).

In a standard pressure or reaction plate, the highest stresses occur at the outer diameter (OD) and inner diameter (ID) edges. The distributed loading by the clutch apply piston causes the plate to deflect to a generally conical shape of some angle theta. With the addition of the annular web 380a, 408a, more material is being strained, resulting in lower stresses and less deflection theta. Thus, the addition of the web 380a, 408a produces a plate having a stiffness comparable to the entire volume from 1D to OD of the reaction plate 380 or pressure plate member 408 having been filled with material.

Additionally, the inside diameter of the annular web portion 380a is fitted closely with the outside diameter of clutch retainer fingers 341 such that the fingers 341 and snap ring 384 are better supported (effectively stronger) against axial piston loading and centrifugally induced loads.

BLEEDER BALL CHECK VALVES

As illustrated in FIGS. 2A and 2B, the reaction shaft member 164 and reaction shaft support 204 are shown. The reaction shaft member 164 is welded at 210 to the reaction shaft support 204. The reaction shaft support 204 comprises a plate portion 232 integral with a hub portion 234. The plate portion 232 includes a pair of circumferentially spaced apertures 236 and 238. Each aperture 236 and 238 has a corresponding passageway 240 communicating therewith and an outlet port 242. Bleeder means such as bleeder ball check valves or dribblers, generally indicated at 244 and 246, are disposed in apertures 236 and 238, respectively, to fill the clutch apply cavities 620 and 622 of the overdrive 304 and reverse 306 assemblies, respectively, as soon as possible after the input clutch retainer 326 begins to rotate and to assure that some fluid always dribbles thereto.

The bleeder ball check valves 244 and 246 each include a screen 248, a ball support 250 disposed in the passageway 240 and a ball 252 supported therein. The ball 252 moves to open and close a narrow aperture or orifice 254 in the ball support 250. The screen 248 acts as a filter to prevent plugging of the orifice 254. The ball support 250 is also formed with inwardly directed fingers 255 which limit the axial movement of the ball 252. The bleeder ball check valves 244 and 246 allow one-way fluid flow to either the overdrive 304 or reverse 306 clutch assemblies, respectively. The size or diameter of the orifice 254 is selected to maintain a minimum pressure, i.e. approximately 0.25 to 2 p.s.i., in the clutch apply cavities 620 and 622 at all times.

In operation, fluid flows from the torque converter 110 to reaction shaft support 204. When either clutch assembly 304, 306 is released, fluid enters apertures 236 or 238 in the plate portion 232 and flows through the corresponding orifice 254 in ball support 250 due to the pressure differential between the fluid pressure from the torque converter 110 and the respective clutch apply cavity being vented. Fluid displaces and moves past the ball 252 to the overdrive 304 or reverse 306 clutch assemblies. When the clutch apply cavity is filled, the fluid pressure moves the ball 252 to close the orifice 254 to prevent backflow. Thus, the bleeder ball check valves 236 and 238 provide fluid to keep the clutch apply cavities 620 and 622, respectively, filled and maintain a pressure balance on the first hydraulic piston 386 whenever rotation exists.

PRESSURE BALANCED PISTONS

Referring to FIG. 1B, the first hydraulic piston 386 includes at least one bleeder orifice 256. The bleeder orifice 256 is typically 0.020 inches in diameter and communicates axially through the first hydraulic piston 386 A filter such as a screen 258 is disposed in the bleeder orifice 256 to prevent plugging of the bleeder orifice 256 by dirt and other contaminants.

In operation the first hydraulic piston 386 is displaced axially by fluid pressure in the clutch apply cavities 622 and 620 for the application of either the reverse 306 or overdrive 304 clutch assemblies, respectively. When that application is removed, the first hydraulic piston 386 must return to its substantially centered or non-applied position. Due to the centrifugal force acting on the rotating fluid in either of the clutch apply cavities 620 or 622 which applied the piston 386, an unbalanced pressure will exist and cause the first hydraulic piston 386 to be biased and remain in that position even though the fluid apply line is vented. The bleeder orifice 256 acts as a means to allow fluid to pass through the first hydraulic piston 386 due to this differential pressure and allows the first hydraulic piston 386 to be centered by the spring 404 since any centrifugal fluid pressure in the clutch apply cavity is balanced by a comparable centrifugal fluid pressure on the opposite side when both clutch apply cavities 620 and 622 are filled. The second hydraulic piston 410 has a similar bleed orifice (not shown) and secondary source of fluid to fill its pressure balance cavity.

DOUBLE-ACTING SPRING

As illustrated in FIGS. 1C, 3A and 3B, a means such as a double-acting spring 404 locates and returns the first hydraulic piston 386. The double-acting spring 404 is a Belleville like spring. The double-acting spring 404 is also annular and conically shaped with circumferentially spaced and inwardly extending fingers 405. The double-acting spring 404 provides the advantage of saving space axially in the transmission 100 due to its compactness. In other words, a conventional coil spring would increase the axial length of the transmission 100 as compared to the double-acting spring 404.

The spring 404 is double-acting; that is, it is applied in two directions at two different focal points. As illustrated in FIG. 3C. when the first hydraulic piston 386 is located or substantially centered in its non-engaged or non-applied position between the input clutch retainer hub 312 and the input clutch retainer 326, the double-acting spring 404 maintains a four point contact. The double-acting spring 404 contacts the snap ring 394, the shoulder portion 313 of the input clutch retainer hub 312, the inner periphery of the first hydraulic piston 386 and one end of the hub portion 328 of the input clutch retainer 326.

When the first hydraulic piston 386 applies the overdrive clutch 304 the double-acting spring 404 is displaced toward the torque converter 110. As illustrated in FIG. 3D, the double-acting spring 404 at its outer periphery contacts the shoulder portion 313 of the input clutch retainer hub 312 and the inner periphery of the first hydraulic piston 386. The double-acting spring 404 applies a return force toward its centered position at the inner periphery of the first hydraulic piston 386.

When the first hydraulic piston 386 applies the reverse clutch 306, the double-acting spring 404 is displaced axially in a direction away from the torque converter 110. As illustrated in FIG. 3E, the double-acting spring 404 contacts the snap ring 394 and the end of the hub portion 328 of the input clutch retainer 326. The double-acting spring 404 applies a return force toward its centered position at the snap ring 394.

In other words, double-acting spring 404 applies a force at its outer periphery in the direction of the torque converter 110 to move the first hydraulic piston 386 axially toward the torque converter 110. This focal point is located at the inner periphery thereof. The double-acting spring 404 also applies a force at its inner periphery in the direction of the output gear 534 to move the first hydraulic piston 386 toward the output gear 534. This focal point is located at the outer periphery of the double-acting spring 404.

Additionally, the double-acting spring 404 is preloaded either by the first hydraulic piston 386 or the input clutch retainer hub 312. Since the double-acting spring 404 usually bends as a cantilever beam, the preloading of the spring 404 by the inner periphery of the first hydraulic piston 386 produces a tip deflection at the outer periphery of the double-acting spring 404, resulting in a gap between the spring 404 and the snap ring 394. Preloading at the shoulder 313 by the input clutch retainer hub 312 produces a tip deflection in the opposite direction, reducing the gap between the double-acting spring 404 and snap ring 394 by a large amount. As a result, the double-acting spring 404 will take some intermediate or statically indeterminate position, distributing the load to all four apply points previously described. In other words, the axial lash in the piston position is removed by the double-acting spring 404 as it deforms to take a statically indeterminate position between the first hydraulic piston 386 and the input clutch retainer hub 312.

The present invention has been described in an illustrative manner. It is to be understood that the terminology which has been used is intended to be in the nature of words of description rather than of limitation.

Obviously, many modifications and variations are possible in light of the above teachings. Therefore, the subject invention may be practiced otherwise than as specifically described. 

What is claimed is:
 1. A vehicle transmission assembly having at least four forward speeds and at least one reverse speed, comprising:a transmission housing; a rotatable input member; an output member; a gear assembly for changing the ratio of torque between said input member and said output member; not more than three clutch assemblies to selectively couple said rotatable input member to predetermined gears of said gear assembly and not more than two brake assemblies to selectively couple predetermined gears of said gear assembly to the transmission housing; a piston member having a non-applied position and being movable axially in one direction to apply a first clutch assembly and movable axially in an opposite direction to apply a second adjacent clutch assembly; said gear assembly comprising a first planetary gearset and a second planetary gearset axially spaced from said first planetary gearset; said first planetary gearset comprising a first sun gear, a first annulus gear and a first planet carrier having a plurality of circumferentially spaced first pinion gears disposed between said first sun gear and said first annulus gear; said second planetary gearset comprising a second sun gear, a second annulus gear and a second planet carrier having a plurality of circumferentially spaced second pinion gears disposed between said second sun gear and said second annulus gear; first means for allowing said first annulus gear and said second planet carrier to rotate at the same time; second means for allowing said first planet carrier and said second annulus gear to rotate at the same time; wherein a first clutch assembly selectively couples said input member to said second sun gear, a second clutch assembly selectively couples said input member to said first planet carrier, a third clutch assembly selectively couples said input member to said first sun gear, a first brake assembly selectively couples said first sun gear to said transmission housing, and a second brake assembly selectively couples said first planet carrier and said second annulus gear to said transmission housing; and an output member connected to said second planet carrier to rotate at the same time.
 2. A vehicle transmission assembly having at least four forward speeds and at least one reverse speed, comprising:a transmission housing; a rotatable input member; an output member; a gear assembly for changing the ratio of torque between said input member ad said output member; not more than three clutch assemblies to selectively couple said rotatable input member to predetermined gears of said gear assembly and not more than two brake assemblies to selectively couple predetermined gears of said gear assembly to the transmission housing; a piston member having a non-applied position and being movable axially in one direction to apply a first clutch assembly and movable axially in an opposite direction to apply a second adjacent clutch assembly; said gear assembly comprising a first planetary gearset and a second planetary gearset axially spaced from said first planetary gearset; said first planetary gearset comprising a first sun gear, a first annulus gear and a first planet carrier having a plurality of circumferentially spaced first pinion gears disposed between said first sun gear and said first annulus gear; said second planetary gearset comprising a second sun gear, a second annulus gear and a second planet carrier having a plurality of circumferentially spaced second pinion gears disposed between said second sun gear and said second annulus gear; first means for allowing said first annulus gear and said second planet carrier to rotate at the same time; second means for allowing said first planet carrier and said second annulus gear to rotate at the same time; and an output member connected to said second planet carrier to allow said output member and said second planet carrier to rotate at the same time.
 3. A vehicle transmission assembly having at least four forward speeds and at least one reverse speed, comprising:a transmission housing; a rotatable input member; an output member; a gear assembly for changing the ratio of torque between said input member and said output member; not more than three clutch assemblies to selectively couple said rotatable input member to predetermined gears of said gear assembly and not more than two brake assemblies to selectively couple predetermined gears of said gear assembly to the transmission housing; a piston member having a non-applied position and being movable axially in one direction to apply a first clutch assembly and movable axially in an opposite direction to apply a second adjacent clutch assembly; said gear assembly comprising a first planetary gearset and a second planetary gearset axially spaced from said first planetary gearset; said first planetary gearset comprising a first sun gear, a first annulus gear and a first planet carrier having a plurality of circumferentially spaced first pinion gears disposed between said first sun gear and said first annulus gear; said second planetary gearset comprising a second sun gear, a second annulus gear and a second planet carrier having a plurality of circumferentially spaced second pinion gears disposed between said second sun gear and said second annulus gear; wherein a first clutch assembly selectively couples said input member to said second sun gear, a second clutch assembly selectively couples said input member to said first planet carrier, a third clutch assembly selectively couples said input member to said first sun gear, a first brake assembly selectively couples said first sun gear to said transmission housing, and a second brake assembly selectively couples said first planet carrier and said second annulus gear to said transmission housing; and an output member connected to said second planet carrier to allow said output member and said second planet carrier to rotate at the same time. 